Spark ignition type internal combustion engine

ABSTRACT

In an internal combustion engine, a variable compression ratio mechanism able to change a mechanical compression ratio and a variable valve timing mechanism able to control the closing timing of an intake valve are provided. At the time of engine low load operation, the mechanical compression is made the maximum so that the expansion ratio becomes 20 or more. Further, at the time of engine low load operation, in part of the operating region or all of the operating region, the actual compression ratio is lowered compared with the time of the engine high load operation and, at the time of engine low load operation, at least when the actual compression ratio is being lowered compared with the time of engine high load operation, the throttle valve is made to close.

TECHNICAL FIELD

The present invention relates to a spark ignition type internalcombustion engine.

BACKGROUND ART

Known in the art is a spark ignition type internal combustion engineprovided with a variable compression ratio mechanism able to change amechanical compression ratio and a variable valve timing mechanism ableto control a closing timing of an intake valve, performing asupercharging action by a supercharger at the time of engine medium loadoperation and engine high load operation, and increasing the mechanicalcompression ratio and delaying the closing timing of the intake valve asthe engine load becomes lower in the state holding the actual combustionratio constant when the engine operation is shifted from the high loadoperation to the medium load operation (for example, see Japanese PatentPublication (A) No. 2004-218522).

However, this document does not allude at all to the actual compressionratio when the engine load is low.

DISCLOSURE OF THE INVENTION

An object of the present invention is to provide a spark ignition typeinternal combustion engine enabling stable combustion to be maintained.

According to the present invention, there is provided a spark ignitiontype internal combustion engine provided with a variable compressionratio mechanism able to change a mechanical compression ratio, avariable valve timing mechanism able to control a closing timing of anintake valve, and a throttle valve arranged in an engine intake passagefor controlling an intake air amount, wherein a mechanical compressionratio is made higher at a time of engine low load operation than at atime of engine high load operation, an actual compression ratio isreduced in part of an operating region or all of the operating region ata time of engine low load operation compared with a time of engine highload operation, and a throttle valve is made to close at least when theactual compression ratio is reduced at a time of engine low loadoperation compared with at a time of engine high load operation.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is an overview of a spark ignition type internal combustionengine.

FIG. 2 is a disassembled perspective view of a variable compressionratio mechanism.

FIG. 3 is a side cross-sectional view of the illustrated internalcombustion engine.

FIG. 4 is a view of a variable valve timing mechanism.

FIG. 5 is a view showing the amounts of lift of the intake valve andexhaust valve.

FIG. 6 is a view for explaining the engine compression ratio, actualcompression ratio, and expansion ratio.

FIG. 7 is a view showing the relationship between the theoreticalthermal efficiency and expansion ratio.

FIG. 8 is a view for explaining an ordinary cycle and superhighexpansion ratio cycle.

FIG. 9 is a view showing the change in mechanical compression ratio etc.in accordance with the engine load.

FIG. 10 is a view showing the change in mechanical compression ratioetc. in accordance with the engine load.

FIG. 11 is a flowchart for operational control.

FIG. 12 is a view showing a map, of the closing timing of the intakevalve etc.

BEST MODE FOR CARRYING OUT THE INVENTION

FIG. 1 shows a side cross-sectional view of a spark ignition typeinternal combustion engine.

Referring to FIG. 1, 1 indicates a crank case, 2 a cylinder block, 3 acylinder head, 4 a piston, 5 a combustion chamber, 6 a spark plugarranged at the top center of the combustion chamber 5, 7 an intakevalve, 8 an intake port, 9 an exhaust valve, and 10 an exhaust port. Theintake port 8 is connected through an intake branch tube 11 to a surgetank 12, while each intake branch tube 11 is provided with a fuelinjector 13 for injecting fuel toward a corresponding intake port 8.Note that each fuel injector 13 may be arranged at each combustionchamber 5 instead of being attached to each intake branch tube 11.

The surge tank 12 is connected through an intake duct 14 to an aircleaner 15, while the intake duct 14 is provided inside it with athrottle valve 17 driven by an actuator 16 and an intake air amountdetector 18 using for example a hot wire. On the other hand, the exhaustport 10 is connected through an exhaust manifold 19 to a catalyticconverter 20 housing for example a three-way catalyst, while the exhaustmanifold 19 is provided inside it with an air-fuel ratio sensor 21.

On the other hand, in the embodiment shown in FIG. 1, the connectingpart of the crank case 1 and the cylinder block 2 is provided with avariable compression ratio mechanism A able to change the relativepositions of the crank case 1 and cylinder block 2 in the cylinder axialdirection so as to change the volume of the combustion chamber 5 whenthe piston 4 is positioned at compression top dead center, and there isfurther provided with an actual compression action start timing changingmechanism B able to change a start timing of an actual compressionaction. Note that in the embodiment shown in FIG. 1, this actualcompression action start timing changing mechanism B is comprised of avariable valve timing mechanism able to control the closing timing ofthe intake valve 7.

The electronic control unit 30 is comprised of a digital computerprovided with components connected with each other through abidirectional bus 31 such as a ROM (read only memory) 32, RAM (randomaccess memory) 33, CPU (microprocessor) 34, input port 35, and outputport 36. The output signal of the intake air amount detector 18 and theoutput signal of the air-fuel ratio sensor 21 are input throughcorresponding AD converters 37 to the input port 35. Further, theaccelerator pedal 40 is connected to a load sensor 41 generating anoutput voltage proportional to the amount of depression L of theaccelerator pedal 40. The output voltage of the load sensor 41 is inputthrough a corresponding AD converter 37 to the input port 35. Further,the input port 35 is connected to a crank angle sensor 42 generating anoutput pulse every time the crankshaft rotates by for example 30°. Onthe other hand, the output port 36 is connected through the drivecircuit 38 to a spark plug 6, fuel injector 13, throttle valve driveactuator 16, variable compression ratio mechanism A, and variable valvetiming mechanism B.

FIG. 2 is a disassembled perspective view of the variable compressionratio mechanism A shown in FIG. 1, while FIG. 3 is a sidecross-sectional view of the illustrated internal combustion engine.Referring to FIG. 2, at the bottom of the two side walls of the cylinderblock 2, a plurality of projecting parts 50 separated from each other bya certain distance are formed. Each projecting part 50 is formed with acircular cross-section cam insertion hole 51. On the other hand, the topsurface of the crank case 1 is formed with a plurality of projectingparts 52 separated from each other by a certain distance and fittingbetween the corresponding projecting parts 50. These projecting parts 52are also formed with circular cross-section cam insertion holes 53.

As shown in FIG. 2, a pair of cam shafts 54, 55 is provided. Each of thecam shafts 54, 55 has circular cams 56 fixed on it able to be rotatablyinserted in the cam insertion holes 51 at every other position. Thesecircular cams 56 are coaxial with the axes of rotation of the cam shafts54, 55. On the other hand, between the circular cams 56, as shown by thehatching in FIG. 3, extend eccentric shafts 57 arranged eccentricallywith respect to the axes of rotation of the cam shafts 54, 55. Eacheccentric shaft 57 has other circular cams 58 rotatably attached to iteccentrically. As shown in FIG. 2, these circular cams 58 are arrangedbetween the circular cams 56. These circular cams 58 are rotatablyinserted in the corresponding cam insertion holes 53.

When the circular cams 56 fastened to the cam shafts 54, 55 are rotatedin opposite directions as shown by the solid line arrows in FIG. 3(A)from the state shown in FIG. 3(A), the eccentric shafts 57 move towardthe bottom center, so the circular cams 58 rotate in the oppositedirections from the circular cams 56 in the cam insertion holes 53 asshown by the broken line arrows in FIG. 3(A). As shown in FIG. 3(B),when the eccentric shafts 57 move toward the bottom center, the centersof the circular cams 58 move to below the eccentric shafts 57.

As will be understood from a comparison of FIG. 3(A) and FIG. 3(B), therelative positions of the crank case 1 and cylinder block 2 aredetermined by the distance between the centers of the circular cams 56and the centers of the circular cams 58. The larger the distance betweenthe centers of the circular cams 56 and the centers of the circular cams58, the further the cylinder block 2 from the crank case 1. If thecylinder block 2 moves away from the crank case 1, the volume of thecombustion chamber 5 when the piston 4 is positioned as compression topdead center increases, therefore by making the cam shafts 54, 55 rotate,the volume of the combustion chamber 5 when the piston 4 is positionedas compression top dead center can be changed.

As shown in FIG. 2, to make the cam shafts 54, 55 rotate in oppositedirections, the shaft of a drive motor 59 is provided with a pair ofworm gears 61, 62 with opposite thread directions. Gears 63, 64 engagingwith these worm gears 61, 62 are fastened to ends of the cam shafts 54,55. In this embodiment, the drive motor 59 may be driven to change thevolume of the combustion chamber 5 when the piston 4 is positioned atcompression top dead center over a broad range. Note that the variablecompression ratio mechanism A shown from FIG. 1 to FIG. 3 shows anexample. Any type of variable compression ratio mechanism may be used.

On the other hand, FIG. 4 shows a variable valve timing mechanism Battached to the end of the cam shaft 70 for driving the intake valve 7in FIG. 1. Referring to FIG. 4, this variable valve timing mechanism Bis provided with a timing pulley 71 rotated by an engine crank shaftthrough a timing belt in the arrow direction, a cylindrical housing 72rotating together with the timing pulley 71, a shaft 73 able to rotatetogether with an intake valve drive cam shaft 70 and rotate relative tothe cylindrical housing 72, a plurality of partitions 74 extending froman inside circumference of the cylindrical housing 72 to an outsidecircumference of the shaft 73, and vanes 75 extending between thepartitions 74 from the outside circumference of the shaft 73 to theinside circumference of the cylindrical housing 72, the two sides of thevanes 75 formed with hydraulic chambers for advancing 76 and usehydraulic chambers for retarding 77.

The feed of working oil to the hydraulic chambers 76, 77 is controlledby a working oil feed control valve 78. This working oil feed controlvalve 78 is provided with hydraulic ports 79, 80 connected to thehydraulic chambers 76, 77, a feed port 82 for working oil dischargedfrom a hydraulic pump 81, a pair of drain ports 83, 84 and a spool valve85 for controlling connection and disconnection of the ports 79, 80, 82,83, 84.

To advance the phase of the cams of the intake valve drive cam shaft 70,in FIG. 4, the spool valve 85 is made to move to the right, working oilfed from the feed port 82 is fed through the hydraulic port 79 to thehydraulic chambers for advancing 76, and working oil in the hydraulicchambers for retarding 77 is drained from the drain port 84. At thistime, the shaft 73 is made to rotate relative to the cylindrical housing72 in the arrow direction.

As opposed to this, to retard the phase of the cams of the intake valvedrive cam shaft 70, in FIG. 4, the spool valve 85 is made to move to theleft, working oil fed from the feed port 82 is fed through the hydraulicport 80 to the hydraulic chambers for retarding 77, and working oil inthe hydraulic chambers for advancing 76 is drained from the drain port83. At this time, the shaft 73 is made to rotate relative to thecylindrical housing 72 in the direction opposite to the arrows.

When the shaft 73 is made to rotate relative to the cylindrical housing72, if the spool valve 85 is returned to the neutral position shown inFIG. 4, the operation for relative rotation of the shaft 73 is ended,and the shaft 73 is held at the relative rotational position at thattime. Therefore, it is possible to use the variable valve timingmechanism B so as to advance or retard the phase of the cams of theintake valve drive cam shaft 70 by exactly the desired amount.

In FIG. 5, the solid line shows when the variable valve timing mechanismB is used to advance the phase of the cams of the intake valve drive camshaft 70 the most, while the broken line shows when it is used to retardthe phase of the cams of the intake valve drive cam shaft 70 the most.Therefore, the opening time of the intake valve 7 can be freely setbetween the range shown by the solid line in FIG. 5 and the range shownby the broken line, therefore the closing timing of the intake valve 7can be set to any crank angle in the range shown by the arrow C in FIG.5.

The variable valve timing mechanism B shown in FIG. 1 and FIG. 4 is oneexample. For example, a variable valve timing mechanism or other varioustypes of variable valve timing mechanisms able to change only theclosing timing of the intake valve while maintaining the opening timingof the intake valve constant can be used.

Next, the meaning of the terms used in the present application will beexplained with reference to FIG. 6. Note that FIGS. 6(A), (B), and (C)show for explanatory purposes an engine with a volume of the combustionchambers of 50 ml and a stroke volume of the piston of 500 ml. In theseFIGS. 6(A), (B), and (C), the combustion chamber volume shows the volumeof the combustion chamber when the piston is at compression top deadcenter.

FIG. 6(A) explains the mechanical compression ratio. The mechanicalcompression ratio is a value determined mechanically from the strokevolume of the piston and combustion chamber volume at the time of acompression stroke. This mechanical compression ratio is expressed by(combustion chamber volume+stroke volume)/combustion chamber volume. Inthe example shown in FIG. 6(A), this mechanical compression ratiobecomes (50 ml+500 ml)/50 ml=11.

FIG. 6(B) explains the actual compression ratio. This actual compressionratio is a value determined from the actual stroke volume of the pistonfrom when the compression action is actually started to when the pistonreaches top dead center and the combustion chamber volume. This actualcompression ratio is expressed by (combustion chamber volume+actualstroke volume)/combustion chamber volume. That is, as shown in FIG.6(B), even if the piston starts to rise in the compression stroke, nocompression action is performed while the intake valve is opened. Theactual compression action is started after the intake valve closes.Therefore, the actual compression ratio is expressed as follows usingthe actual stroke volume. In the example shown in FIG. 6(B), the actualcompression ratio becomes (50 ml+450 ml)/50 ml=10.

FIG. 6(C) explains the expansion ratio. The expansion ratio is a valuedetermined from the stroke volume of the piston at the time of anexpansion stroke and the combustion chamber volume. This expansion ratiois expressed by the (combustion chamber volume+stroke volume)/combustionchamber volume. In the example shown in FIG. 6(C), this expansion ratiobecomes (50 ml+500 ml)/50 ml=11.

Next, the most basic features of the present invention will be explainedwith reference to FIG. 7 and FIG. 8. Note that FIG. 7 shows therelationship between the theoretical thermal efficiency and theexpansion ratio, while FIG. 8 shows a comparison between the ordinarycycle and superhigh expansion ratio cycle used selectively in accordancewith the load in the present invention.

FIG. 8(A) shows the ordinary cycle when the intake valve closes near thebottom dead center and the compression action by the piston is startedfrom near substantially compression bottom dead center. In the exampleshown in this FIG. 8(A) as well, in the same way as the examples shownin FIGS. 6(A), (B), and (C), the combustion chamber volume is made 50ml, and the stroke volume of the piston is made 500 ml. As will beunderstood from FIG. 8(A), in an ordinary cycle, the mechanicalcompression ratio is (50 ml+500 ml)/50 ml=11, the actual compressionratio is also about 11, and the expansion ratio also becomes (50 ml+500ml)/50 ml=11. That is, in an ordinary internal combustion engine, themechanical compression ratio and actual compression ratio and theexpansion ratio become substantially equal.

The solid line in FIG. 7 shows the change in the theoretical thermalefficiency in the case where the actual compression ratio and expansionratio are substantially equal, that is, in the ordinary cycle. In thiscase, it is learned that the larger the expansion ratio, that is, thehigher the actual compression ratio, the higher the theoretical thermalefficiency. Therefore, in an ordinary cycle, to raise the theoreticalthermal efficiency, the actual compression ratio should be made higher.However, due to the restrictions on the occurrence of knocking at thetime of engine high load operation, the actual compression ratio canonly be raised even at the maximum to about 12, accordingly, in anordinary cycle, the theoretical thermal efficiency cannot be madesufficiently high.

On the other hand, under this situation, the inventors strictlydifferentiated between the mechanical compression ratio and actualcompression ratio and studied the theoretical thermal efficiency and asa result discovered that in the theoretical thermal efficiency, theexpansion ratio is dominant, and the theoretical thermal efficiency isnot affected much at all by the actual compression ratio. That is, ifraising the actual compression ratio, the explosive force rises, butcompression requires a large energy, accordingly even if raising theactual compression ratio, the theoretical thermal efficiency will notrise much at all.

As opposed to this, if increasing the expansion ratio, the longer theperiod during which a force acts pressing down the piston at the time ofthe expansion stroke, the longer the time that the piston gives arotational force to the crankshaft. Therefore, the larger the expansionratio is made, the higher the theoretical thermal efficiency becomes.The broken line ε=10 in FIG. 7 shows the theoretical thermal efficiencyin the case of fixing the actual compression ratio at 10 and raising theexpansion ratio in that state. In this way, it is learned that theamount of rise of the theoretical thermal efficiency when raising theexpansion ratio in the state where the actual compression ratio ismaintained at a low value and the amount of rise of the theoreticalthermal efficiency in the case where the actual compression ratio isincreased along with the expansion ratio as shown by the solid line ofFIG. 7 will not differ that much.

If the actual compression ratio is maintained at a low value in thisway, knocking will not occur, therefore if raising the expansion ratioin the state where the actual compression ratio is maintained at a lowvalue, the occurrence of knocking can be prevented and the theoreticalthermal efficiency can be greatly raised. FIG. 8(B) shows an example ofthe case when using the variable compression ratio mechanism A andvariable valve timing mechanism B to maintain the actual compressionratio at a low value and raise the expansion ratio.

Referring to FIG. 8(B), in this example, the variable compression ratiomechanism A is used to lower the combustion chamber volume from 50 ml to20 ml. On the other hand, the variable valve timing mechanism B is usedto delay the closing timing of the intake valve until the actual strokevolume of the piston changes from 500 ml to 200 ml. As a result, in thisexample, the actual compression ratio becomes (20 ml+200 ml)/20 ml=11and the expansion ratio becomes (20 ml+500 ml)/20 ml=26. In the ordinarycycle shown in FIG. 8(A), as explained above, the actual compressionratio is about 11 and the expansion ratio is 11. Compared with thiscase, in the case shown in FIG. 8(B), it is learned that only theexpansion ratio is raised to 26. This is the reason that it is calledthe “superhigh expansion ratio cycle”.

In an internal combustion engine, generally speaking, the lower theengine load, the worse the thermal efficiency, therefore to improve thethermal efficiency at the time of engine operation, that is, to improvethe fuel consumption, it becomes necessary to improve the thermalefficiency at the time of engine low load operation. On the other hand,in the superhigh expansion ratio cycle shown in FIG. 8(B), the actualstroke volume of the piston at the time of the compression stroke ismade smaller, so the amount of intake air which can be sucked into thecombustion chamber 5 becomes smaller, therefore this superhigh expansionratio cycle can only be employed when the engine load is relatively low.Therefore, in the present invention, at the time of engine low loadoperation, the superhigh expansion ratio cycle shown in FIG. 8(B) isset, while at the time of engine high load operation, the ordinary cycleshown in FIG. 8(A) is set.

Next, the operational control as a whole will be explained withreference to FIG. 9.

FIG. 9 shows the changes in the mechanical compression ratio, expansionratio, closing timing of the intake valve 7, actual compression ratio,the amount of intake air, opening degree of the throttle valve 17, andpumping loss along with the engine load under a certain engine speed.Note that in the embodiment according to the present invention,ordinarily the average air-fuel ratio in the combustion chamber 5 isfeedback controlled to the stoichiometric air-fuel ratio based on theoutput signal of the air-fuel ratio sensor 21 so that the three-waycatalyst in the catalytic converter 20 can simultaneously reduce theunburned HC, CO, and NO_(x) in the exhaust gas.

Now, as explained above, at the time of engine high load operation, theordinary cycle shown in FIG. 8(A) is executed. Therefore, as shown inFIG. 9, at this time, since the mechanical compression ratio is madelow, the expansion ratio becomes low. As shown by the solid line in lowin FIG. 9, the closing timing of the intake valve 7 is advanced as shownby the solid line in FIG. 5. Further, at this time, the amount of intakeair is large. At this time, the opening degree of the throttle valve 17is maintained fully opened or substantially fully opened, so the pumpingloss becomes zero.

On the other hand, as shown by the solid line in FIG. 9, when the engineload becomes lower, the closing timing of the intake valve 7 is retardedso as to reduce the intake air amount along with that. Further, at thistime, the mechanical compression ratio is increased as the engine loadbecomes lower as shown in FIG. 9 so that the actual compression ratio ismaintained substantially constant. Therefore, the expansion ratio isalso increased as the engine load becomes lower. Note that at this timeas well, the throttle valve 17 is held in the fully open orsubstantially fully open state. Therefore, the intake air amount fedinto the combustion chamber 5 is controlled by changing the closingtiming of the intake valve 7 without relying on the throttle valve 17.At this time as well, the pumping loss becomes zero.

In this way, when the engine load becomes lower from the engine highload operation state, the mechanical compression rate is made toincrease along with the reduction in the intake air amount under asubstantially constant actual compression ratio. That is, the volume ofthe combustion chamber 5 when the piston 4 reaches compression top deadcenter is reduced proportionally to the reduction in the intake airamount. Therefore, the volume of the combustion chamber 5 when thepiston 4 reaches compression top dead center changes in proportion tothe intake air amount. Note that the air-fuel ratio in the combustionchamber 5 at this time is made the stoichiometric air-fuel ratio, so thevolume of the combustion chamber 5 when the piston 4 reaches compressiontop dead center changes in proportion to the fuel amount.

If the engine load becomes further lower, the mechanical compressionratio is further made to increase. When the engine load falls to themedium load L₁ closer to low load, the mechanical compression ratioreaches the limit mechanical compression ratio constituting thestructural limit of the combustion chamber 5. In the region of a loadlower than the engine load L₁ where the mechanical compression ratioreaches the limit mechanical compression ratio, the mechanicalcompression ratio is held at the limit mechanical compression ratio.Therefore, at the time of low load side engine medium load operation andat the time of engine low load operation, the mechanical compressionratio becomes maximum and the expansion ratio also becomes maximum. Inother words, at the time of low load side engine medium load operationand at the time of engine low load operation, the mechanical compressionratio is made maximum so that the maximum expansion ratio is obtained.

On the other hand, in the embodiment shown in FIG. 9, even when theengine load becomes lower than L₁, as shown by the solid line in FIG. 9,the closing timing of the intake valve 7 is retarded as the engine loadbecomes lower. When the engine load falls to L₂, the closing timing ofthe intake valve 7 becomes the limit closing timing where the intake airamount fed into the combustion chamber 5 can be controlled. When theclosing timing of the intake valve 7 reaches the limit closing timing,in the region of a load lower than the engine load L₂ when the closingtiming of the intake valve 7 reaches the limit closing timing, theclosing timing of the intake valve 7 is held at the limit closingtiming.

When the closing timing of the intake valve 7 is held at the limitclosing timing, the intake air amount soon no longer can be controlledby changing the closing timing of the intake valve 7. In the embodimentshown in FIG. 9, at this time, that is, in the region of a load lowerthan the engine load L₂ when the closing timing of the intake valve 7reaches the limit closing timing, the intake air amount fed into thecombustion chamber 5 is controlled by the throttle valve 17. However, ifthe intake air amount is controlled by the throttle valve 17, thepumping loss increases as shown in FIG. 9.

On the other hand, as shown in FIG. 9, when the engine load is higherthan L₁, that is, at the time of the high load side engine medium loadoperation and at the time of engine high load operation, the actualcompression ratio is maintained at substantially the same actualcompression ratio for the same engine speed. As opposed to this, whenthe engine load is lower than L₂, that is, when the mechanicalcompression ratio is held at the limit mechanical compression ratio, theactual compression ratio is determined by the closing timing of theintake valve 7. If the closing timing of the intake valve 7 is retardedas in the engine load between L₁ and L₂, the actual compression ratiofalls. If the closing timing of the intake valve 7 is held at the limitclosing timing as in the engine load lower than L₂, the actualcompression ratio is maintained constant.

Therefore, in the embodiment shown in FIG. 9, at the time of the lowload side engine medium load operation and at the time of engine lowload operation where the engine load is lower than L₁, the actualcompression ratio is reduced compared with the actual compression ratioat the time of engine high load operation. In this regard, if the actualcompression ratio falls in this way, the temperature in the combustionchamber 5 at the compression end falls and ignition and combustion ofthe fuel deteriorate. However, in this case, if closing the throttlevalve 17, the throttling action of the intake air flow by the throttlevalve 17 causes disturbances inside the combustion chamber 5 and therebycan improve the ignition and combustion of the fuel.

On the other hand, in the operating region where the engine load islower than L₂, the operation can also be controlled by increasing theair-fuel ratio the lower the engine load in the state with the throttlevalve 17 held fully open. However, in the present invention, to improvethe ignition and combustion of the fuel, as shown in FIG. 9, theoperation is controlled by closing the throttle valve 17 when the actualcompression ratio falls.

In the embodiment shown in FIG. 10, even when the engine load becomeslower than L₁, while higher than L₃, the closing period of the intakevalve 7 is held constant and during this time the throttle valve 17 ismade to close. On the other hand, when the engine load becomes lowerthan L₃ as well, the closing timing of the intake valve 7 is retardeduntil L₂, that is, until the closing timing of the intake valve 7reaches the limit closing timing. That is, in this embodiment, while theclosing timing of the intake valve 7 is being retarded along with thefall in the engine load, even if the engine load changes, there are loadregions L₃ to L₁ where the closing timing of the intake valve 7 cannotbe changed. In these load regions L₃ to L₁, the intake air amount iscontrolled by the throttle valve 17.

In this regard, in this embodiment, in the region of a load higher thanthe load regions L₃ to L₁, the throttle valve 17 is held in the fullyopen state. In the operating region where the engine load is lower thanL₃, the actual compression ratio is reduced compared with the time ofengine high load operation. In this embodiment as well, the throttlevalve 17 is made to close when the actual compression ratio is reduced.

Therefore, in the present invention, at the time of engine low loadoperation, the actual compression ratio is reduced in part of theoperating region or all of the operating region compared with the timeof engine high load operation, and, at the time of engine low loadoperation, at least when the actual compression ratio is reducedcompared with the time of engine high load operation, the throttle valve17 is made to close.

Explained more specifically, in the present invention, in the operatingregion where the load is lower than the engine load L₁ when themechanical compression ratio reaches the limit mechanical compressionratio, in part or all of the operating region, the actual compressionratio is lowered compared with the operating region where the load ishigher than this operating region. In this way, the throttle valve 17 isclosed when the actual compression ratio is reduced in this way.

Note that, in the embodiment of the present invention, as will beunderstood from FIG. 9 and FIG. 10, in the operating region where theload is lower than the engine load L₁ when the mechanical compressionratio reaches the limit mechanical compression ratio, the intake airamount is controlled by controlling one of the closing timing of theintake valve 7 and the opening timing of the throttle valve 17.

As explained above, in the super high expansion ratio cycle shown inFIG. 8(B), the expansion ratio is made 26. The higher this expansionratio the better, but as will be understood from FIG. 7, it is possibleto obtain a considerably high theoretical thermal efficiency if 20 ormore even for the practically usable limit actual compression ratio ε=5.Therefore, in the present invention, the variable compression ratiomechanism A is formed so that the expansion ratio becomes 20 or more.

On the other hand, as shown by the broken line in FIG. 9, it is possibleto control the intake air amount without regard to the throttle valve 17by advancing the closing timing of the intake valve 7 as the engine loadbecomes lower. Therefore, expressing this so that both the case shown bythe solid line in FIG. 9 and the case shown by the broken line arecovered, in the embodiment of the present invention, the closing timingof the intake valve 7 is made to shift as the engine load becomes lowerin the direction away from the intake bottom dead center BDC until thelimit closing timing L₂ enabling control of the intake air amount fedinto the combustion chamber.

FIG. 11 shows the operational control routine. Referring to FIG. 11,first, at step 100, the target actual compression ratio is calculated.Next, at step 101, the closing timing IC of the intake valve 7 iscalculated from the map shown in FIG. 12(A). That is, the closing timingIC of the intake valve 7 required for feeding the required intake airamount into the combustion chamber 5 is stored as a function of theengine load L and engine speed N in the form of the map as shown in FIG.12(A) in advance in the ROM 32. The closing timing IC of the intakevalve 7 is calculated from this map.

Next, at step 102, the mechanical compression ratio CR is calculated.Next, at step 103, the opening degree of the throttle valve 17 iscalculated. The opening degree θ of this throttle valve 17 is stored asa function of the engine load L and engine speed N in the form of a mapas shown in FIG. 12(B) in advance in the ROM 32. Next, at step 104, thevariable compression ratio mechanism A is controlled so that themechanical compression ratio becomes the mechanical compression ratioCR, the variable valve timing mechanism B is controlled so that theclosing timing of the intake valve 7 becomes the closing timing IC, andthe throttle valve 17 is controlled so that the opening degree of thethrottle valve 17 becomes the opening degree θ.

LIST OF REFERENCE NOTATIONS

-   1 . . . crank case-   2 . . . cylinder block-   3 . . . cylinder head-   4 . . . piston-   5 . . . combustion chamber-   7 . . . intake valve-   70 . . . intake valve drive cam shaft-   A . . . variable compression ratio mechanism-   B . . . variable valve timing mechanism

1. A spark ignition type internal combustion engine provided with avariable compression ratio mechanism able to change a mechanicalcompression ratio, a variable valve timing mechanism able to control aclosing timing of an intake valve, and a throttle valve arranged in anengine intake passage for controlling an intake air amount, wherein amechanical compression ratio is made higher at a time of engine low loadoperation than at a time of engine high load operation, an actualcompression ratio is reduced in part of an operating region or all ofthe operating region at a time of engine low load operation comparedwith a time of engine high load operation, and a throttle valve is madeto close at least when the actual compression ratio is reduced at a timeof engine low load operation compared with at a time of engine high loadoperation.
 2. A spark ignition type internal combustion engine asclaimed in claim 1, wherein at the time of engine low load operation,the mechanical compression ratio is made a maximum mechanicalcompression ratio.
 3. A spark ignition type internal combustion engineas claimed in claim 1, wherein at the time of engine low load operation,the expansion ratio is made 20 or more.
 4. A spark ignition typeinternal combustion engine as claimed in claim 1, wherein saidmechanical compression ratio is increased to a limit mechanicalcompression ratio as the engine load becomes lower and the mechanicalcompression ratio is held at said limit mechanical compression ratio ina region of a load lower than the engine load when said mechanicalcompression ratio reaches said limit mechanical compression ratio.
 5. Aspark ignition type internal combustion engine as claimed in claim 4,wherein in an operating area of a load lower than the engine load whensaid mechanical compression ratio reaches said limit mechanicalcompression ratio, in part or all of said operating area, the actualcompression ratio is reduced compared with the operating region of aload higher than said operating area.
 6. A spark ignition type internalcombustion engine as claimed in claim 4, wherein in an operating area ofa load lower than the engine load when said mechanical compression ratioreaches said limit mechanical compression ratio, the intake air amountis controlled by controlling one of the closing timing of the intakevalve and an opening degree of the throttle valve.
 7. A spark ignitiontype internal combustion engine as claimed in claim 1, wherein at thetime of engine high load operation, the throttle valve is held at thefully open state.
 8. A spark ignition type internal combustion engine asclaimed in claim 1, wherein the closing timing of the intake valve isshifted as the engine load becomes lower in a direction away from intakebottom dead center until the limit closing timing enabling control ofthe intake air amount fed into the combustion chamber.
 9. A sparkignition type internal combustion engine as claimed in claim 8, whereinin the region of a load higher than the engine load when the closingtiming of the intake valve reaches said limit closing timing, thethrottle valve is held at the fully open state.
 10. A spark ignitiontype internal combustion engine as claimed in claim 8, wherein in theregion of a load lower than the engine load when the closing timing ofthe intake valve reaches said limit closing timing, the intake airamount is controlled by the throttle valve.
 11. A spark ignition typeinternal combustion engine as claimed in claim 8, wherein while theclosing timing of the intake valve is being shifted as the engine loadbecomes lower in a direction away from intake bottom dead center, thereare load regions where the closing timing of the intake valve cannot bechanged even if the engine load changes, and, in said load regions, theintake air amount is controlled by the throttle valve.
 12. A sparkignition type internal combustion engine as claimed in claim 11, whereinthe throttle valve is held in a fully open state in a region of a loadhigher than said load region.